The present invention relates to a variable-displacement vane pump comprising a main body having a cavity, in which is movable a ring containing a rotor rotating about a fixed axis. The rotor has a number of vanes, one end of each of which rests on the inner surface of the ring during rotation.

Means are also provided which, depending on a control pressure, move the ring between a centred position with respect to the rotation axis of the rotor, in which no pumping action takes place, and a predetermined eccentric position with respect to the rotation axis of the rotor.

Vane pumps of the above type are currently used to pump various fluids, such as oil in an internal combustion engine.

As is known, in pumps of the above type, at high rotation speeds of the shaft connected to the rotor, the gaps between adjacent vanes on the pump fail to fill completely, thus resulting in forces impairing operation of the pump.

To counteract the increase in such forces, counteracting springs are traditionally used, but are extremely rigid and therefore do not deform easily.

FR 2 195 271 (PEUGEOT & RENAULT) discloses a pumping system comprising a variable-delivery vane pump; the system also comprising a user device (not shown) connected to the pump by a delivery conduit, and pressure regulating means for adjusting the eccentricity of a ring with respect to a rotor, both forming part of the pump, so that the pump assumes a balanced configuration such as to supply the oil flow requested by said user device; and wherein the pressure regulating means comprise hydraulic dissipating means for imparting to the oil inside a regulating chamber forming part of the pump a pressure lower than a control pressure. However, in such a document the dissipating means are located in series with the control device. The position of the dissipating means with regard to the discharging valve does not permit a fine control of the pressure in the control chamber.

In DE 32 14 212 (TEVES) a pressure control device for vane cell pumps is described. Such a control device has a servo piston cylinder unit and a control piston cylinder unit, the cylinder chambers of which are connected to a consumer pressure line of the pump and the pistons of which act on opposite sides of a servo element influencing the pump delivery, the control piston having a larger piston surface than the servo piston and being spring-loaded into the position of maximum delivery of the servo element, and a pressure limiting valve which is arranged in a line leading to an unpressurised vessel. In order to simplify the construction of the pressure control device and in order to reduce the leakage points without resorting to spring-loaded valve slides, it is proposed that in the pressure line section leading to the control piston cylinder unit a diaphragm be arranged and that the pressure limiting valve be connected between the diaphragm and the control piston cylinder unit.

In such a document the diaphragm constitutes the dissipating means. However, the pressure limiting valve does not feel the pressure in the delivery conduit. Therefore, the pressure control into the control piston cylinder unit is not precise.

In WO 02/052155 (INAGUMA YOSHIHARU et al.) a vane pump is described. Such pump has a cam ring mounted in an adapter and made movable in the radial directions, and an internal pressure and a loading pressure, which are introduced into action chambers formed on the two sides of the cam ring, between the upstream and downstream of a variable orifice are controlled by a differential pressure control valve to control the discharge flow rate in response to the pump rotation speed. The differential pressure control valve is activated by the internal pressure and the loading pressure, which are introduced into action chambers formed on the two end sides, and by a valve pushing spring for biasing the differential pressure control valve to the internal pressure action chamber, and this biasing force by the valve pushing spring is increased/decreased according to the increase/decrease in the loading pressure. This increase/decrease is caused, for example, by a loading pressure responding piston which is biased by a piston pushing spring so that its leading end comes into the internal pressure action chamber and abuts against the differential pressure control valve. However, in such a document no precise pressure control is achieved.

It is therefore an object of the present invention to provide for hydraulic control of a variable-delivery vane pump, particularly at high speed.

According to the present invention, there is provided a pumping system employing a variable-displacement vane pump, as claimed in Claim 1.

A number of non-limiting embodiments of the present invention will be described by way of example with reference to the accompanying drawings, in which:

• Figure 1 shows a first embodiment of the present invention;
• Figure 2 shows a first configuration of a second embodiment;
• Figure 3 shows a second configuration of the second embodiment in Figure 2;
• Figure 4 shows a first configuration of a detail of the second embodiment in Figures 2 and 3;
• Figure 5 shows a second configuration of the Figure 4 detail;

Number 10 in Figure 1 indicates a variable-delivery vane pump forming part of a pumping system 100 in accordance with the present invention.

Pump 10 comprises, in known manner, a main body 11 having a cavity 12 in which a ring 13 translates as described in detail later on.

Ring 13 houses a rotor 14 having a number of vanes 15, which move radially inside respective radial slits 16 formed in rotor 14, which is rotated in the direction indicated by arrow W (see below).

Main body 11 is closed by a cover not shown in the drawings.

In known manner, rotor 14 houses a shaft 17 connected mechanically to rotor 14; and a floating ring 18 surrounding shaft 17, and on which the other ends of vanes 15 rest.

Shaft 17 therefore has a centre P1 which is fixed at all times; and ring 13 has a centre P2.

The distance P1P2 represents the eccentricity E of pump 10.

As is known, by varying eccentricity E, the delivery of pump 10 can be varied as required by a user device UT downstream from pump 10 (see below).

User device UT may be defined, for example, by an internal combustion engine (not shown).

As shown in Figure 1, ring 13 has a projection 19 housed partly in a chamber 20; and a projection 21 housed partly in a chamber 22. Projections 19 and 21 are located on opposite sides of centre P2 of ring 13, and have respective front surfaces A1 and A2 facing chambers 20 and 22 respectively. For reasons explained in detail later on, surface A2 is larger than surface A1. More specifically, tests and calculations have shown surface A2 must be 1.4 to 1.7 times larger than surface A1.

A spring 22a inside chamber 22 exerts a small force on surface A2 to restore the system to a condition of maximum eccentricity E when system 100 is idle.

In the Figure 1 embodiment, chambers 20 and 22 are formed in main body 11 of pump 10.

Main body 11 also comprises an intake port 23 for drawing oil from a tank 24; and a delivery port 25 for feeding oil to user device UT.

A delivery conduit 26 extends from delivery port 25 to feed user device UT.

As shown in Figure 1, a first portion of the oil supplied to user device UT is diverted to chamber 20 by a conduit 27, and a second portion of the oil is fed to chamber 22 by a conduit 28.

More specifically, the second portion of the oil in conduit 28 is almost all fed to chamber 22 by a conduit 28a via a dissipating device 29, in which a calibrated pressure loss occurs as the oil actually flows through it.

Conduit 28 is connected to a valve 30 by a conduit 28b.

Valve 30 comprises a cylinder 31 housing a piston 32.

More specifically, as shown in Figure 1, piston 32 comprises a first portion 32a and a second portion 32b connected to each other by a rod 32c.

Portions 32a and 32b are equal in cross section to cylinder 31, whereas rod 32c is smaller in cross section than cylinder 31.

Cylinder 31 has a port 33 connected hydraulically to chamber 22 by a conduit 34.

Conduit 28b substantially provides for picking up a delivery pressure signal in conduit 28, so as to act on the front surface A3 of portion 32a of piston 32.

The dash line in Figure 1 shows the situation in which port 33 is closed by second portion 32b.

As described in more detail later on, as the delivery pressure (p1) increases alongside an increase in the operating speed of pump 10, greater force is exerted on surface A3 and, on reaching the preload value of a spring 36, moves piston 32 to allow oil flow from conduit 34 through port 33 and along a conduit 35 into tank 24.

At the start of conduit 35, the oil is at atmospheric pressure (po).

Piston 32 is stressed elastically by a suitably sized spring 36 designed to generate a force which only permits movement of piston 32 when the delivery pressure (p1) on surface A3 reaches a given value.

A return conduit 37 from user device UT to tank 24 completes pumping system 100.

In the known art, eccentricity E is normally regulated by diverting a portion of the oil supply into a chamber, in which the delivery pressure acts directly on the ring; and an elastic counteracting force, generated by a spring, acts on the opposite side of the ring, so that the pump is set to an eccentricity E value ensuring the oil pressure and flow requested by the user device.

As stated, however, at high rotation speed of the shaft, and therefore of the rotor and vanes, the gaps between adjacent vanes fail to fill completely, thus resulting in undesired forces which, in addition to high rotation speed of the rotor, also depend on the temperature and chemical-physical characteristics of the oil.

Incomplete fill has the side effect of generating a force acting in the direction indicated by arrow F1 in Figure 1.

Consequently, the user device fails to obtain the required delivery pressure, on account of this undesired force which, as stated, is substantially caused by incomplete oil fill of the gaps between the vanes.

By way of a solution to the problem, an attempt has been made to disassociate control from the above negative internal forces by providing for so-called "hydraulic control".

With reference to Figure 1, since, as stated, surface A2 is larger (preferably 1.4 to 1.7 times larger) than surface A1, it follows that, if the delivery pressure (p1) were present in both chambers 20 and 22, a force would be generated in the direction indicated by arrow F2 to compensate the force produced by incomplete fill of gaps 15a, thus resulting in maximum eccentricity E.

In this case, however, there would be no adjustment. To achieve the desired adjustment, therefore, the oil pressure in chamber 22 must be made lower than in chamber 20.

Consequently, when the delivery pressure (p1) reaches a value capable of generating sufficient force on surface A3 of portion 32a to overcome the elastic force of spring 36, piston 32 moves into the configuration shown by the continuous line in Figure 1, in which rod 32c of piston 32 is located at port 33, and so permits oil flow from chamber 22 to conduit 34, and along conduit 35 into tank 24.

Oil therefore also flows along conduit 28a and through dissipating device 29, so that, as opposed to the delivery pressure (p1), a lower pressure (p2) is present in chamber 22.

In other words, the pressure (p2) in chamber 22 is lower than the pressure (p1) in chamber 20, thus disassociating the two pressures to enable ring 13 to move in the direction indicated by arrow F1 to establish a balanced eccentricity E value producing the desired oil flow to user deice UT.

More specifically, as delivery pressure increases to a value (p*) determined by the characteristics of spring 36, piston 32 begins moving so that part of the oil leaks through port 33. In other words, valve 30 also acts as a pressure dissipating device to assist in creating the desired pressure (p2) in chamber 22.

(p1) and (p*) are equal at the end of the transient state

The system has also proved stable.

That is, adjustment continues for as long as permitted by piston 32, i.e. control is taken over by valve 30, which is regulated solely by delivery pressure (p1) and totally unaffected by undesired internal forces.

Conversely, in known regulating systems, when delivery pressure (p1) increases, it remains constant for a while, and then decreases.

With the system employed in system 100, on the other hand, on reaching the value required by user device UT, pressure (p1) remains constant, even at extremely high rotation speeds of rotor 14.

When delivery pressure reaches pressure value (p*), substantially determined by the characteristics of spring 36, generation of pressure (p2) commences, and ring 13 begins moving in the direction of arrow F1 to reduce eccentricity E and therefore the displacement of pump 10. Consequently, delivery pressure falls, and tends to assume a value below (p*), so that piston 32 moves into an intermediate balance position reducing the size of port 33.

Displacement remains fixed up to a given pressure value and, as pump speed increases, tends to increase delivery. When a given pressure value (p*) is reached, valve 30 opens, and oil flows along conduit 34, through port 33, and along conduit 35 to tank 24, so that the pressure (p2) in chamber 22 is lower than (p1), and ring 13 moves in the direction of arrow F1 to reduce displacement and therefore oil flow to user device UT.

In a second embodiment shown in Figures 2 to 5, dissipating device 29 and valve 30 are replaced by a three-way slide valve 50.

Valve 50 comprises a cylinder 51 housing a slide 52 stressed by a spring 53.

As shown more clearly in Figures 4 and 5, slide 52 comprises a first portion 52a, a second portion 52b, and a third portion 52c. Portions 52a and 52b are connected by a rod 52d, and portions 52b and 52c are connected by a rod 52e.

Cylinder 51 comprises four ports 54, 55, 56, 57. More specifically, port 54 defines the first way of three-way valve 50, ports 56 and 57 together define the second way, and port 55 defines the third way.

Slide 52 is controlled by delivery pressure (p1).

As shown in Figure 4, the value of ε1, which represents the size of port 56, must be greater than ε2, i.e. the size of the closed area covered by portion 52b of slide 52.

As shown in Figures 2 and 4, when delivery pressure (p1) is below a given value (p*), the oil pressure in chamber 22 assumes the delivery pressure (p1) value, pump 10 is set to maximum eccentricity, and no pumping action occurs.

As delivery pressure (p1) increases, slide 52 begins moving in the direction of arrow F3, so that oil flows from port 54 to port 56, and from port 57 to port 55 and into tank 24 maintained at atmospheric pressure (p0) (Figures 3, 5).

In other words, oil leaks from valve 50, thus resulting in a load loss, so that the pressure (p2) of the oil in chamber 22 assumes an intermediate value between delivery pressure (p1) and the atmospheric pressure (p0) of tank 24.